GB1567331A

GB1567331A - Planetary gesring

Info

Publication number: GB1567331A
Application number: GB41743/76A
Authority: GB
Original assignee: Balcke Duerr AG
Current assignee: Balcke Duerr AG
Priority date: 1975-10-11
Filing date: 1976-10-07
Publication date: 1980-05-14
Also published as: CH626692A5; NL7611141A; FR2327453A1; RO71333B; IS2343A7; CA1060234A; SE434878B; JPS6112137B2; IT1066678B; FI762769A; RO71333A; DD128028A5; AR209996A1; PT65597B; NO763431L; AU1855476A; ES452298A1; LU75964A1; JPS5247164A; BR7606755A

Description

PATENT SPECIFICATION ( 11) 1567 331

t 91 ( 21) Application No 41743/76 ( 22) Filed 7 Oct 1976 f ( 31) Convention Application No 2545681 ( 19) ( 32) Filed 11 Oct 1975 ( 31) Convention Application No 2551083 ( 32) Filed 14 Nov 1975 ( 31) Convention Application No 2617951 y ( 32) Filed 24 April 1976 in ( 33) Federal Republic of Germany (DE) ( 44) Complete Specification published 14 May 1980 ( 51) INT CL 3 F 16 H 1/32 ( 52) Index at acceptance F 2 Q 7 A 3 F 7 H 4 B 7 H 5 C ( 54) IMPROVEMENTS IN OR RELATING TO PLANETARY GEARING ( 71) We, BALCKE-DORR AKTIENGESELLSCHAFT, a German Body Corporate, of Homberger Strasse 2, D-4030 Ratingen, German Federal Republic, do hereby declare the invention, for which we pray that a patent may be granted to us, and the method by which it is to be performed, to be particularly described in

and by the following statement: 5

This invention relates to planetary gearing.

When two central gears having different numbers of teeth are arranged so that the teeth of one of the gears are juxtaposed with the teeth of the other of the gears, at least two "virtual tooth rows" are formed by the teeth of the one gear overlapping the teeth of the other gear, each tooth gap of each virtual tooth row 10 being formed by a side of one of the teeth of the one gear -and a side of one of the teeth of the other gear, as hereinafter explained.

Such planetary gearing has become known heretofore from German Patent DT-PS 929,771 In this heretofore known planetary gearing, a maximum of two 1 H 1 teeth of a planet gear meshes with the virtual tooth row The involute toothing 15 E conventional with planetary gearing effects a rolling on one another of the teeth that are in engagement and thereby the transmission of forces only in one contacting line, respectively In accordance with the invention, on the other hand, almost all of the teeth of the planet gear mesh with virtual tooth gaps and are in areally or flatly adjacent engagement in order: 20 a) to be able to transmit greater torque for the same dimensions of the gearing or b) to be able to construct smaller gearing for the same transmitted torque.

The invention provides planetary gearing comprising a toothed planet gear, first and second relatively rotatable central gears having juxtaposed rows of teeth 25 meshing with said toothed planet gear, said juxtaposed rows of teeth having different numbers of teeth and forming at least two virtual tooth rows by overlapping, each tooth gap of each virtual tooth row being formed by a side of one of the teeth of the first central gear and a side of one of the teeth of the second central gear and each virtual tooth row defining a curve passing through the roots 30 of the tooth gaps of that virtual tooth row and a cam rotatable about an axis and having a cam outline engaging the planet gear for guiding and driving the planet gear, all of the teeth being of substantially triangular cross-section and having substantially flat sides, the sides of the teeth of the planet gear substantially flatly engaging the sides of the teeth of one of the virtual tooth rows, the pitch of each 35 meshing planet gear tooth being substantially equal to the pitch of the respective meshing tooth of said virtual tooth row, and the centroid of the area enclosed by the cam outline coinciding with the centroid of the area enclosed by the curve passing through the roots of the tooth gaps of said one virtual tooth row.

By way of example five embodiments of planetary gearing and modifications 40 thereto according to the invention will now be described with reference to the accompanying drawings, in which:Figure IA is a diagram illustrating how the virtual tooth rows provided by the planetary gearing according to the invention are formed; Figure 1 is a radial sectional view of Figure 32 taken along the line I-I in the 45 k i 1 2,/ I,J_ 67 I direction of the arrows showing an embodiment of the planetary gearing according to the invention wherein two of the internally toothed central gears are shown having different numbers of teeth, and employing only one virtual tooth row; Figure 2 is a fragmentary radial sectional view similar to that of Figure 1 showing another embodiment of the planetary gearing with two of the externally 5 toothed central gears having different numbers of teeth; Figure 3 is a diagrammatic, radial quarter sectional view of the periphery of a planetary gearing with two central gears having a tooth number difference of four:

Figure 3 A is an enlargement of the portion of Figure 3 indicated at I Ila in Figure 3; 10 Figures 4 to 7 are various diagrammatic views of details of the planetary gearing showing the disposition of the teeth of the planet gear between tooth sides of the virtual tooth row; Figures 8 to i 5 are schematic views and plot diagrams, where clearly applicable, explaining the deviation of the curve passing through the roots of the 15 tooth gaps of the virtual tooth row from the "smooth" form and the rounding off in the respective kink or jog locations; Figure 16 is a schematic radial sectional view of a gearing with a tooth number difference of two and employing a pair of virtual tooth rows; Figure 17 A is a half longitudinal sectional view of part of a gearing according 20 to the invention; Figure 17 is table of values which, together with Figure 17, serves for explaining the reduction ratio; Figure 18 a partial radial sectional view of a planet gear according to the invention showing the tooth row thereof in the form of a zig-zag metal sheet; 25 Figure 19 is a perspective view of part of the zig-zag metal sheet of Figure 18; Figures 20, 21 and 22 are respective perspective, sectional and top plan views of the shiftable disposition of teeth on the planet gear; Figures 23 to 29 are sectional views of various constructions of the teeth of the planet gear; 30 Figure 30 is a quarter radial sectional view of a planetary gearing according to the invention; Figure 31 is a fragmentary enlarged plan view of the bearing band formed with a slit; Figure 32 is a longitudinal sectional view of the planetary gearing of Figure 1; 35 Figure 33 is a radial sectional view of a gearing similar to that of Figure 1; Figure 34 is a longitudinal sectional view of Figure 35 taken along the line XXXIV-XXXIV in the direction of the arrows; Figure 35 is a radial sectional view of Figure 34 taken along the line XXXVXXXV in the direction of the arrows; and 40 Figure 36 is a longitudinal sectional view of a control gearing constructed in accordance with the invention.

Figure IA shows portions of two internally toothed relatively rotatable central gears having different numbers of teeth The teeth of one of the gears having tooth sides or flanks 1, 2 are juxtaposed with the teeth of the other of the gears having 45 tooth sides or flanks 3, 4 The effect of the teeth of the one gear overlapping the teeth of the other gear is to define two series of virtual teeth Each tooth gap of one of the series is formed by adjacent sides 1, 3 of said one and said other gears, respectively, and each tooth gap of the other series is formed by adjacent sides 2, 4 of said one and said other gears, respectively Each series is a "virtual tooth row" 50 Referring to Figures 1 and 32 of the drawings, there is shown therein a planetary gearing constructed in accordance with the invention and having two internally toothed relatively rotatable central gears 42 and 44 According to the cross-sectional view shown in Figure 1, the internally toothed central gear 42 is disposed behind the internally toothed central gear 44, as viewed into the plane of 55 the figure For this reason, the flanks of the teeth of the central gear 42 are covered in part by the teeth of the forward central gear 44 and, where visible, are illustrated in broken lines in Figure 1.

In the right-hand upper quadrant of Figure 1, only the teeth 46 of the internally toothed central gear 44 are shown It is apparent in the view of Figure 1 that the zig 60 zag lines formed by the flanks of the tooth rows of the internally toothed central gears 42 and 44 overlap in such manner that two virtual tooth rows are formed Of the two tooth virtual rows, the "utilized" row is accentuated because the outer teeth 48 of a planet gear 50, shown stippled in Figure 1, engage therein or mesh therewith It is noted especially in the left hand side of Figure 1, that yet a second 65 I567 11 3 1,567,331 3 virtual tooth row is formed which is not utilized in Figure 1 This second virtual tooth row corresponds identically with the first-mentioned virtual toothed row, except that it is offset, however, by a given angle Instead of the firstmentioned utilized virtual toothed row of Figure 1, the second virtual tooth row therein could ff 51 be utilized, only the rotary sense or direction thereof being changed Hereinafter, 5 discussion is had only with respect to a virtual tooth row which is utilized The teeth of all of the gears have a triangular cross-section and substantially flat flanks.

The roots 52 and 54 of the internally toothed central gears 42 and 44 are disposed on a root circle 56 having a center 58 which is the point of intersection of the common central axis 60 of both internally toothed central gears 42, 44, as 10 viewed into the plane of the figure The roots 62 of the tooth gaps of the virtual tooth row, contrarily, lie close to a curve 64 substantially having the shape of a circle which is hereinafter referred to as "virtual root circle" or quite generally as "virtual root curve" The center 66 of the virtual root circle 64 is the point of 1 S intersection of the orbiting axis 68 into the plane of the drawings of Figure 1 The 15 orbiting axis 68 is offset from the center 58 of root circle 56 During rotation of the planet gear 50, the center 66 (the orbiting axis 68) describes a circle about the center 58 (the central axis 60).

It is also apparent in Figure 1 that the outer or external teeth 48 of the planet 201 gear 50 engage in the "tooth gap" of the virtual tooth row in a manner that the 20 points or tips of the outer teeth 48 of the planet gear 50 press forward up to the roots 62 of the virtual tooth row The height of the outer teeth 48 of the planet gear is about double the spacing between the centers 58 and 66, and therefore the number of teeth 48 of the planet gear 50 which are in free transmitting engagement 25: with the virtual tooth row is substantially equal to the total number of teeth of the 25 planet gear less the difference between the number of teeth of the pair of central gears.

A cam disc 70 is disposed within the planet gear 50; rollers 72 are provided as force-transmitting bearings between the cam disc 70 and the planet gear 50 in order 3 t J to facilitate the rotation of the planet gear 50 with respect to the cam disc 70 30 The rearward ceintirafgear 42 of the illustrated gearing, as viewed into the plane of Figure 1, has seventy-eight teeth 74, whereas the forward central gear 44 possesses slightly more, namely eighty teeth 46 The externally toothed planet gear carries seventy-nine teeth 48, the width of which (perpendicular to the plane of the drawing) is great according to Figures 6 and 7 that they mesh with both 35 internally toothed central gears 42 and 44 For the sake of clarity, less teeth are shown in Figure 1 According to Figure 32, the internally toothed central gear 42 is held stationary; in this case, only a drive of the cam disc 70 by the drive shaft 76, having an axis coinciding with the central axis 60, is involved The internally 4 Q go toothed central gear 44 is connected with the driven shaft 78 The outline of the 40 cam disc 70 has a centre coinciding with the orbiting axis 68, and therefore rotation of the cam disc 70 about the central axis 60 effects a rotation of the planet gear 50, the outer teeth 48 of which are braced in the teeth 74 of the firmly held internally toothed central gear 42 From the aforementioned number of teeth (seventyeight and eighty) of the central gears 42 and 44, respectively, there is given, 45 independently of the number of teeth of the planet gear 50, from the equation a) either a reduction ratio of 39, which means that for thirty-nine revolutions of the drive shaft 76, the drive shaft 78 makes one revolution, i b) or a reduction ratio of 40 with reversed rotary sense if the other central gear 50 is held stationary.

Figure 2 shows a gearing according to the invention with a tooth count difference of two Two externally toothed central gears 82 and 84 are surrounded by an internally toothed planet gear 90 which is, in turn surrounded by a hollow cam disc 86, the central gear 82 being disposed behind the central gear 84 as viewed in 55 direction into the plane of the drawing of Figure 2 The tooth rows of the externally toothed central gears 82 and 84 overlap to form two virtual tooth rows, with one of which the inner or internal teeth 88 of the planet gear 90 mesh In the left-hand part of Figure 2, the internal teeth 88 of the planet gear are formed from a zig-zag metal sheet which is shown in perspective view in Figure 19 In the right-hand side of 60 Figure 2, on the other hand, another embodiment of the internal teeth 88 is shown.

i j 1 t h 1 1 k Obviously, in a planet gear, only an embodiment of the internal teeth will be used.

During rotation of the hollow cylindrical cam disc 86 about the central axis 60, the internal teeth 88 press into the virtual tooth row due to the centre of the outline of the cam 86 coinciding with the orbiting axis 68 and thus effect mutual relative rotation of the central gears 82 and 84 It is apparent that a gearing according to 5 Figure 2 with externally toothed central gears 82 and 84, that are surrounded by the internally tooth planet gear 90, is constructed and functions, in principle, in the same manner as a gearing with internally toothed central gears 42 and 44 which surround an externally toothed planet gear 50 as in Figure 1.

Therefore, to explain and describe the planetary gearing of the invention 10 hereinafter, a gearing with internally toothed central gears and at least one externally toothed planet gear according to Figure 1 are used The explanations and descriptions obviously apply also to planetary gearing that are constructed in accordance with Figure 2.

Figure 3 shows schematically in section a quadrant of a gearing according to 15 the invention The difference in the number of teeth on the central gears is 4 The gearing of Figure 3 is constructed substantially like the gearing according to Figure 1 and accordingly possesses two internally toothed relative rotatable central gears 42 and 44 Regions of the internally toothed central gear 42 are covered by the central gear 44 and therefore represented by a partially broken zig-zag line The 20 central gear 44 is not covered and is therefore represented by a solid zig-zag line.

Both rows of teeth of the internally toothed central gears 42 and 44 form two virtual tooth rows The external teeth 48 of the planet gear mesh with or engage in the virtual tooth row.

The flank angles and the spacings of the virtual tooth gaps vary over the 25 periphery of the virtual tooth row In planetary gearing with two central gears as shown in Figure 3, if ar=half the flank angle of the tooth gaps of the used virtual tooth rows, ac=half the flank angle of the teeth of the planet gear, (T,-a)m=half the difference between the flank angle of the used virtual tooth 30 row, on the one hand, and the flank angle of the planet gear, on the other hand, at a location m, i 1 =the number of teeth of a first central gear, Z 2 =the number of teeth of a second central gear, m=the number (ordinal number) of the tooth under consideration, as counted 35 from a location at which a,,-a-= 0 (note Figure 3), Tmax,=the maximal pitch (spacing of the tooth gaps) of the virtual tooth row, Tmn,=the minimal pitch (spacing of the tooth gaps) of the virtual tooth row, then both of the following equations are valid:

(oc>-oc)= 3 03 ( 4 1) m ( 1) 40 Tmin v = 1 Tmax v 1 tan ( 90-o C) ( 2) Z 1 +Z 2 The tooth row of the planet gear has given flank angles 2 a and a given pitch T (spacing of the teeth), which is, for example, the mean value between Tmax, and Tmn,v From the equations (l) and ( 2), it is inferred that the deviation or variation of the tooth flank angle of the virtual tooth row from the tooth flank angle of the 45 planet gear as well as the deviation or variation of the pitch or spacing of the virtual tooth row (varying between Tmax, v and Tmin v) from the (constant) pitch or spacing of the planet gear become all the greater, the greater the difference there is between the numbers of teeth, and Z 2 of the central gears For this reason, the difference between the numbers of teeth of the central gears lies sensibly between one and six 50 Higher differences in the numbers of teeth would lead to deviations or variations which would not be controllable at reasonable engineering costs.

A favorable value for the difference in the numbers of teeth is A 9 = 2, as for the central gears of Figures 1 and 2 For medium and high reduction ratios (for example, over 30), the curve passing through the roots of the tooth gaps of the 55 virtual tooth row is, from a practical standpoint, nearly a circle having a centroid I 1,567,331 located eccentric to the central axis of the central gears by about half the height of a tooth of the planet gear If two virtual tooth rows are used, the curve is formed of two circles that have been shifted away from one another.

The feature that the teeth of the planet gear, on both sides thereof, areally or flatly engage the flanks of the used virtual tooth row is realizable in various ways: 5 a) through a constant pitch or spacing (within the tolerances) of the virtual tooth row; this constant pitch or spacing can be largely attained for high reduction ratios, the flank angles of the teeth of the central gears and of the planet gear as well as the diameters thereof being suitably selected; or b) for non-uniform pitch or spacing of the virtual tooth row through peripheral 10 and/or radially shiftable device and/or elastic formation or development of the teeth of the planet gear; this means no elastic formation or development of the planet gear per se.

For high reduction ratios, for example over 40, and for a small difference A i between the numbers of teeth of the central gears, the root circles connecting the 15 roots of the tooth gaps of the central gears to one another coincide For a lower reduction ratio, however, such as under 40, for example, and for a greater difference Ah (greater than 2) between the numbers of teeth of the central gears, the teeth of the central gears differ in the height thereof to an extent that a common root circle would result in much too great a difference Tmax v-Tm In v In this case, it 20 is advantageous, in order to achieve a practically constant pitch or spacing of the virtual tooth row, to dispose the tooth rows of the central gears so that they are halved in the level of a circle i e the central gears have different diameters.

Thereby, for smaller reduction ratios, the difference Tmax v-Tmin v is reduced to such an extent that it lies within the limits of manufacturing tolerances, and individual 25 teeth of the planet gear are disengaged from the virtual tooth row.

For medium and high reduction ratios (for example, over 30), the curve passing through the roots of the tooth gaps of the virtual tooth row is capable of being described with sufficient accuracy by a circle or an otherwise closed curve path or trend line, the tangents of which from point to point of the curve path 30 continuously change the direction thereof For low reduction ratios, somewhat between 10 and 30, it has been found, however, that said curve through such l"smooth" or "jog-free" or "kink-free" curves can no longer be approximated with sufficient accuracy On the contrary, said curves, as explained in greater detail hereinafter, are formed of circular segments, which are either mutually connected 35 by straight lines into which the circular segments run, or merge into one another in intersecting points (inflection points or curve breaking points) wherein the tangents to the curve path unsteadily vary the direction thereof.

For reduction ratios between 10 and 30, special problems occur which, for higher reduction ratios are adjustable yet with relatively simple means, such as 40 elasticity of material, for example, or lie within the frame or limits of the tolerances These special problems which occur for reduction ratios between 10 and 30 are explained hereinafter:

During the rotary movement of the cam, half of the teeth of the planet gear move radially outwardly whereas the other half of the teeth of the planet gear move 45 radially inwardly If complete engagement of the teeth of the planet gear on the flanks of the virtual tooth row is always to be assured, a) the teeth moving radially outwardly must have identically the same velocity +v, and b) the teeth moving radially inwardly must have identically the same velocity 50 The radial movement of the teeth should also take place with constant velocity (+v or -v) i e without acceleration.

It has furthermore been shown that, within the range of lower reduction ratios, the tooth gaps or spaces of the virtual tooth row have such variable flank angles 55 (note Equation 1 and spaces or divisions (note Equation 2) that, with relatively simple means (such as elasticity of material or the hereinafter described free mobility of the teeth of the planet gear, for example) the sought-after flat engagement of the teeth of the planet gear on the flanks of the virtual row of teeth 6 t) cannot be assured Finally, for lower reduction ratios, the hereinaforedescribed 60 deviations of the said curve from an ideal "smooth" curve with which the direction of the tangents over the periphery continuously varies Such deviations could also occur, for example, if the tooth flanks of the central gears were not flat.

Due to such deviations of the virtual tooth row with respect to the tooth row of the planet gear in the spacing or division and in the flank angle as well as due to 65 i G J 1 d I 1,567,331 deviations of the curve passing through the roots of the tooth gaps of the virtual tooth row from an ideal form, the accurate flat engagement of the teeth of the planet gear at the flanks of the virtual tooth row is impaired For higher reduction ratios, these deviations lie within the frame or limits of manufacturing tolerances and of the flexibility of the material, and can therefore practically remain 5 disregarded.

A further feature of the planetary gearing is that the cam has a contour or outline which, in mathematical sense, is similar to the virtual tooth gap curve, wherein locations, at which the direction of the tangents vary unsteadily or irregularly, are rounded, and wherein, between the said curve and the contour or 10 outline of the cam, the planet gear, the teeth thereof as well as the bearing (such as rollers, for example), which transmits the forces thereof, are disposed, and in that the teeth of the planet gear are independently shiftable radially from one another in peripheral direction and/or radially and/or are elastically deformable and possess a variable flank angle which matches the flank angle of the virtual tooth row 15 Due to the rounding of the contour or outline of the cam at those locations thereof at which said curve has locations of unsteadiness in the tangential direction, exactly at these locations the respective tooth of the planet gear thereat is not forced into the virtual tooth row Since, precisely this tooth experiences the greatest acceleration from +v to -v, by uncoupling this tooth out of the virtual 20 tooth row, the disturbing effect thereof upon the smooth course of movement is eliminated This uncoupling is also effected through the hereinaforementioned variable diameter of the central gears.

The periodic variation in the spacing or divisions of the virtual tooth row over the periphery that is revealed in Equation 2 is balanced or equalized in that the 25 teeth of the planet gear are independently shiftable relative to one another in peripheral direction and/or radially and/or are elastically deformable The difference Tmax v-Tmin v is thereby a measure of the required peripheral mobility of the teeth of the planet gear.

Periodic variations of the flank angle of the virtual tooth row are balanced or 30 equalized by a variable flank angle of the teeth of the planet gear.

The tooth moving in the region of the rounded-off locations is reversed in the radial movement thereof In order that it does not thereby disturb the rotary motion, there is provided, in accordance with an added feature of the invention, of the internally toothed central gears, that gear with the greatest number of teeth 35 has the smallest root circle i e the circle that connects the roots of the teeth while, of the externally toothed central gears, that gear with the greatest number of teeth has the greatest root circle Thereby, the tooth of the planet gear located in the region of the rounded-off location is held only by the tooth flanks of the central gear with the greatest number of teeth and does not promote the torque 40 transmission between the central gears It is sufficient to ensure a mobility or pivotability of the teeth and a variability of the flank angle in order to assure the flat engagement of the teeth of the planet gear against the flanks of the virtual tooth row.

As a favorable value of the difference in the number of teeth of the central 45 gears, the number two was mentioned hereinbefore in relation to Figures 1 and 2.

To achieve lower reductions, desirable constructional possibilities arise from the tooth-number difference Ai A= 4 Basically, the same reduction can be effected with the tooth-number difference Ag= 4, as in Figure 3, as with the toothnumber t Xg= 2, when the number of teeth are doubled; for the same diameter of the gearing teeth 50 that are half as high are then obtained.

The curve passing through the roots of the tooth gaps of the virtual tooth row is composed of circular arcs about a center of a circle and is calculable, if the tooth flanks of the central gears are flat, from an equation of the following type:

rv= r sin (A g) As ( 4) 55 in which:

r,=distance of one tooth gap root (at the location m) of the virtual tooth row from the central axis 60 of the central gears, r=radius of the root circle of the central gears with respect to the central axis 60, 60 I 1,567,331 cl Aip=angular difference (peripheral spacing) of the adjacent tooth gap roots of the central gears at the location m, calculated from Equation (I) and a,-a=AT ( 3) 2 a,=flank angle of the tooth space of the virtual tooth row at the location m, S according to Equation ( 1), 5 As=spacing of the straight connecting lines between the tooth space points called (Ap) of the central gears from the foot circle 56, note "Detail 5254 in Figure 3 ".

If the tooth numbers of both central gears differ by A A= 4, four centers M,, M 2, M 3 and M 4 of circles are obtained, the center of which is located on the central axis 10 of both central gears If the tooth numbers of both central gears differ by Ai A= 2, we then have three centers M,, M, and M 3 of circles, in accordance with Equation ( 4) and must distinguish two instances:

a) The teeth of the planet gear mesh with only one virtual tooth row, as is shown in Figure 1 and will be shown in Figure 12 The common center of the three 15 centers M,, M 2 and M 3 of circles is located eccentrically to the central axis 60 of both central gears.

b) The teeth of the planet gear mesh (Figure 16) with both virtual tooth rows as will be shown in Figure 16 whereby respectively less than half of the teeth of the planet gear meshes with the one or the other virtual tooth row The common center 20 of mass of the four centers of circles (M 1 and M 3 of the one virtual tooth row and M, and M 3 of the other virtual tooth row) is located on the central axis 60 of both central gears.

If the common center of mass of the centers of circles is located on the central axis of the central gears, for example AA= 4; or A i= 2 when using a pair of virtual 25tooth rows i e the immediately hereinaforegoing case (b), thus the cam is advantageously formed of two halves with a circular outline or contour adjustable relative to one another, the contour of each half extending, respectively, over somewhat less than a half circle, the center of the circular outline or contour being located in the center of mass in that one of the point pairs M,-M 2 or M 3-M 4 30 adjacent one another, which is more remote from the respective outline or contour.

The bipartite construction of the cam permits the production of the individual parts with relative slight precision because, due to the adjustment of the parts, any inaccuracies that may exist can be equalized or compensated for during 3 i 15 installation It is also possible, during the adjustment, to achieve a given matching 35 or accommodation of the teeth of the planet gear to the virtual tooth row and to effect a subsequent adjustment, as soon as signs of wear appear.

According to Equation ( 4), virtual tooth gap root curves that are to be calculated also materialize for higher reductions, but agree, however, practically with simpler curves, such as circles, for example, within the manufacturing 40 tolerances Within the range of lower reduction ratios discussed herein, the more accurate curve form of the Equation ( 4) is to be taken into account through the use of a bipartite cam.

Referring specifically to Figure 3, the gaps or spaces of the virtual tooth row which is utilized are indicated by ordinal numbers m, counting from a location m= 0 45 at which a a, (note: Equation 1) A single rigid tooth 48 with a flank angle 2 a is illustrated in a tooth gap or space in Figure 3; it is noted that the flank angle 2 a, of the virtual tooth row is greater than 2 ac If such a tooth 48 were illustrated in each tooth gap or space, it would be apparent that the smaller m is, the smaller the so difference is, and when m= 0, the difference has completely vanished If a tooth 50 having a flank angle 2 a that is elastically variable is used instead of the rigid tooth 48, an optimal adjustment or matching of the tooth row of the planet gear to the virtual tooth row is attained Such adjustable or matchable teeth are shown in Figures 2, 18 and 30 as well as Figure 33.

The peripheral angular difference (AT)m between the location m of the virtual 55 tooth gaps or spaces and the location m of the associated tooth of the planet gear is equal to half the angular deviation of the flanks at the same location m as derived from Equation ( 1):

(A 6 s)mi(iv n)m O In this regard it is noted that: 60 i I 1,567,331 ( 3) vv v=av a 8 1,567,331 8 wherein q-angular distance of the vertical tooth bisector of one tooth of the planet gear from the location m= 0, and lv=angular distance of the vertical tooth bisector of the corresponding tooth of the vertical tooth row from the location m= 0 5 The virtual tooth gap root curve 64 of the utilized virtual tooth row is a circular segment having a center M 4 which lies in the quadrant at the upper righthand side of Figure 3 In the right side, non-illustrated quadrant of the gearing, this curve is a circular segment of the same radius from the center M, of the circle which lies in the upper, left-hand side quadrant The respective circle centers M 2 and M 3 (note: 10 the description with respect to Figure 13) are shown for the nonillustrated lower half of the gearing.

The circle centers M 1, M 2, M 3 and M 4 have equal spacing from the centroid of the area enclosed by the contour or outline of the planet gear and from the centroid of the area enclosed by the contour or outline of the cam disc, both centroids being 15 simultaneously the point of intersection of the central axis 60 of both central gears 42 and 44 through the plane of the drawing of Figure 3 The spacing of each individual point of the virtual tooth gap root curve 64 from the centroid at the location m is calculated according to the previously mentioned equation:

rv=r-h sin (A l 1 As ( 4) 20 wherein:

r,=distance of a tooth gap root 62 (at the location m) of the virtual tooth row from the centroid, r=radius of the root circle 56 of the central gear, A)-angular difference (peripheral spacing of the tooth roots 52 and 54 of the 25 central gear at the location m, calculated in accordance with Equations ( 1) and ( 3), 2 act=flank angle of the tooth of the virtual tooth row at the location m (in Figure 3 at m= 4), As=the spacing of the straight connecting line, located between the tooth gap 30 point 52 and 54 of the central gear and identified as (Ap), from the foot circle 56 (note: "Detail 52-54 in Figure 3 A).

The following relationship also holds for the angle p, which is enclosed by the radius r, and a radius r 4 (the spacing of the tooth gap point 62 from the center point M 4) about the circle center M 4 (in Figure 3 or M, or M 2 or M 3) of the respective 35 circular section of the virtual foot gap root curve 64:

I 2 pv = arc sin 2 (+ | sinocv (z f 1 -sin MV) ( 5) The virtual tooth gap root curve is able to be calculated from Equation ( 5) equally as well as from Equation ( 4).

i-k is also, at the respective location m under consideration, the angle between 40 the tangent to the root circle 56, on the one hand, and the virtual tooth gap root curve 64, on the other hand Therefore, p, is also referred to as "inlet angle" or "outlet angle" with which the virtual tooth row runs into the tooth rows or runs out therefrom.

The outline or contour 96 of the cam disc 70 is similar in a mathematical sense 45 to the virtual tooth gap root curve 64 i e the spacing thereof from the virtual tooth gap root curve 64 is constant The left-hand side contour line 96/4, which is associated with the circle center M 4, intersects at the kink or inflection point 98 with the right-hand side contour line 96/1 having the point M, as its center point.

Due to the intersection of the circular contour lines 96/4 and 96/1 at the inflection 50 point 98, there is formed thereat a point of the contour lines wherein the direction of the tangents to the contour 96 varies non-uniformly This is the point wherein the contour or outline is rounded off so that the tooth 48 of the planet gear present at this location is not pressed into the virtual tooth row but, rather, is movable in radial direction The rounded portion 100 indicated in phantom extends advantageously over several virtual tooth gaps.

Figure 4, which shows a detail of Figure 3, illustrates diagrammatically and schematically, the kinematic principle, based upon which, the gearing of the invention of the instant application operates 5 The tooth 48 of the planet gear engages, on the one hand, the flank of the tooth 46 of the forward central gear 44 and, on the other hand, the respective flank of the tooth 74 of the rearward central gear 42, as viewed into the plane of the drawing of Figure 4 Both tooth rows form one with the other the virtual tooth row, the tooth gap point or root 62 of which coincides with the point of the tooth 48, as long as the 10 latter point is not flattened or rounded off in conventional manner.

The teeth 46 and 74 act upon the tooth 48 with forces indicated by the arrows 102 and 104 These forces are broken down into peripheral force components 106 and 108 and into a radial force component 110 It is evident that the peripheral components 106 and 108 cancel each other out with the result that no forces act 15 upon the tooth 48 in peripheral direction On the one hand, this produces or effects the automatic blocking action and, on the other hand has as a consequence thereof that the planet gear is not required to overcome any forces in peripheral direction thereof and can therefore be given a rather thin or narrow, elastic and interrupted construction or can be provided with individual teeth shiftable in peripheral 20 direction without thereby impairing or interfering with the transmissibility of torque to the central gears and with the durability or life-span thereof Only the radial force component 110 acts upon the tooth 48 and presses it against the planet gear These radial force components are, for example, transmitted, over the rollers 72 to the cam disc 70 and cancelled due to the oppositely directed force in 25 accordance with the arrow 114.

Since the planet gear has to absorb similar forces from all the teeth 48 and therewith from all radially directions all round, these forces are extensively cancelled, so that the drive shaft of the cam disc 70 is not stressed in bending, and the structural components (central gears, planet gear) are centered one within the 30 other.

Figure 5 diagrammatically and schematically illustrates a detail of Figure 3 in the vicinity of the locations m= 4 and m= 5 The teeth 74 marked with crosses and belonging to the rearward central gear are partly covered by the teeth 46 of the ^ 3 g forward central gear as viewed in a direction into the plane of the drawing Two 35 diagrammatically represented teeth 48 of the planet gear engage in both virtual tooth gaps or spaces m= 4 and m= 5 It is apparent, initially, that an exact meshing and an exact mutual engagement of the flanks is possible only if both teeth 48 are pivotable independently of one another and are variable in elevation or height.

lid Both teeth 48 of the planet gear are shown, on the one hand, in the solid-line 40 position and, on the other hand, in a phantom-line position In the latter position, they are inserted so far into the virtual tooth row as would correspond substantially to the positions m= 0 and m=l of Figure 3 The spacing of the teeth 48 in the solidline position is the division T, whereas the division in the phantom-line position is the division T, The ratio of the maximal division Tmax to the minimal division 45 Tmh, v is given by Equation ( 2) and is a measure for the required peripheral mobility of the teeth 48 of the planet gear.

During the movement from the solid-line to the phantom-line positions, the tooth 48 slides along the flanks of the teeth 46 and 74 and, like a wedge, forces these teeth apart With a conventional planetary gearing, on the other hand, the tooth of 50 the planet gear rolls off on a single tooth flank of a central gear.

In Figure 6, a tooth 48 of the planet gear is shown diagrammatically in perspective This tooth 48 meshes with the indicated tooth flanks 116 and 118 of the respective teeth 46 and 74 of the central gears.

Figure 7 is a plan view of Figure 6, areas of the tooth flanks 116 and 118 which 55 flatly engage the tooth 48 being marked with little crosses.

Figures 8 to 15 show the direction of the virtual tooth gap root curve from a "smooth" form and the necessity for the rounding off at the kinks or inflection points.

In Figures 8 and 9 as well as Figures 12 to 15, the root circle 56 of the central 60 gears is set with points 0, 1, 2, 3, 4, 5 and 6 distributed equiangularly thereon In order to avoid obliterating details of the figures, only a single common root circle 56 for two internally toothed central gears is shown The possibility of having a pair of separated root circles 561 and 562 for two internally toothed central gears is l W; I 1,567,331 apparent from Figures 30 and 33 The center point 58 of the root circle 56 lies on the central axis 60 of the central gears.

In Figure 8, a curve path 641 in the form of an inner circle is furthermore indicated, which represents the virtual tooth gap root curve under the (not strictly correct) assumption that this virtual tooth gap root curve is exactly circular 5 Figure 10 shows, starting from the point 0, the spacing between the root circle 56 and the curve (or circle) path 641 for the points 1, 2, 3, 4, 5 and 6 These spacings follow a sine curve Since the contour of the cam disc 70, 86 of a planetary gearing in mathematical sense must be constructed similar to the virtual tooth gap root curve, the radial velocity of the individual teeth differs when the angular velocity of 10 the revolving cam disc is constant Neglecting or ignoring the teeth at the points 0 and 6, however, it is thus assumed that the teeth located thereat are out of engagement or unmeshed from the virtual tooth row, so that, as shown in Figure 10, the curve path between the points 1 and 5 are represented practically by a straight line The spacings between the root circle 56 and the curve (or circle) path 641 vary 15 in the region between the points 1 and 5, thus nearly proportionally to the peripheral angle ( For a constant angular velocity of the revolving cam disc, practically constant radial velocity of the individual teeth, which is desirable, is produced in this case.

Figure 9 shows two inner circles having respective center points 122 and 124 20 These circles are connected at the right-hand and the left-hand sides thereof by straight lines 126, the lengths of which are equal to the spacing between the center points 122 and 124 A respective upper and lower half of the circles forms, together with both straight lines 126, a closed curve path 641 in such manner that tangents applied thereto continuously vary the direction thereof In the case of higher 25 reduction ratios, for a tooth count difference of four, a cam disc similar to the curve path 641 of the virtual tooth gap root curve is practically utilizable.

Figure 11 is a view associated with Figure 9 in a similar manner as Figure 10 is associated with Figure 8.

For lower reductions, in the range between i= 10 and i= 30, the curve paths 641 30 shown in Figures 8 and 9 must be replaced by the curve paths 641 shown in Figures 12 and 13, which is given by Equation ( 4) In both Figures 12 and 13, the center point 58 is the point of intersection of the central axis of both central gears through the plane of the drawing.

Using Equation ( 4), with a tooth count difference of the central gears of two, 35 the three circle centers M 1, M 2 and M 3 shown in Figure 12 are obtained The circle center M 2 is the center of a circular segment which extends over barely the lower half of the curve path 641 The circle center M 1 disposed in the quadrant at the upper right-hand side of Figure 12 is the center of a circular segment which lies substantially in the quadrant of the curve path 641 located at the upper left-hand 40 side of Figure 12 In contrast thereto, the circle center M 3 lying in the quadrant at the upper left-hand side of Figure 12 is the center of the circular segment which lies in the quadrant at the upper right-hand side of that figure Both circular segments, which are associated with the circle centers M 1 and M 3, intersect in the upper kink or inflection point 98 and merge without any kink or inflection point into the lower 45 circular segment which is associated with the circle center M 2 This curve path is shown in broken lines in Figure 14 together with the solid-line curve (or circle) path 641 of Figure 8 The deviation is noted which, for a selected reduction ratio of about 6, can definitely play a role, and as well for higher reduction ratios up to 30 The connections in accordance with the invention of the instant application are 50 significant within this range, because of that, the curve path of the virtual tooth gap root curve is so close to a circle, that it is sufficient, for the most part, to provide the cam disc of the planetary gearing with a circular contour or outline.

The kink or inflection point 98 is that point of Figure 3, wherein the roundingoff portion 100 is applied (note: Figure 3) 55 The curve path 641 according to Figure 12 is, roughly speaking, somewhat pear-shaped i e somewhat wider at the bottom than at the top Figure 13 shows the formation of the curve path 641 for the tooth count difference of four In the quadrant at the upper left-hand side of Figure 13, the curve path 641 is a circular segment with circle center M 4 which lies in the quadrant at the upper right-hand 60 side of Figure 13 In the quadrant at the upper right-hand side of Figure 13, the curve path 641 is a circular segment having a circle center M 1 which lies in the quadrant at the upper left-hand side of the figure In the quadrant at the lower right-hand side of Figure 13, the curve path 641 is a circular segment having a circle center M 2 lying in the quadrant at the lower left-hand side of the figure In the 65 lo 1,567,331 quadrant at the lower left-hand side of Figure 13, the curve path 641 is a circular segment having a circle center M 3 lying in the quadrant at the lower right-hand side of the figure The segments which are associated with the circle center points M 3 and M 2 intersect in a lower kink or inflection point 98 The segments which are associated with the circle center points M 1 and M 4 intersect in an upper kink or 5 inflection point 98 The segments which are associated with the circle center points M 4 and M 3 are connected by a short straight line 130, the length of which is equal to the spacing between the circle centers M, and M 2.

Figure 15, similar to Figure 14, shows a comparison of a simplified curve path 641/9 (in solid lines) of the virtual tooth gap root curve, according to Figure 9, with 10 the complex curve path 641/13 (in broken lines) according to Figure 13 It is noted that the deviations increase with increasing tooth count difference.

The "pear-shaped" curve path 641 of the virtual tooth gap root curve results in a gearing as shown in Figure 1 and with which the revolving axis of the planet gear lies eccentrically to the central axis 60 of the central gears This results in a loading 15 or stressing of the drive shaft due to a bending moment If this bending load or stress is to be avoided, the construction of Figure 16 is selected.

From the curve path 641 of the virtual root curve shown in Figure 12 the upper section which is associated with the circle centers M, and M 3 is used From the non-illustrated tooth gap root curve of the other virtual tooth row, which is 20 offset 1800 from the first-mentioned virtual tooth row, that nonillustrated curve portion which is associated with the non-illustrated circle centers of this second virtual tooth row is used Thus, the same as in Figure 13, four circle centers and, therewith, a curve path composed of four circular arcs are obtained By using both virtual tooth rows, in accordance with Figure 16, a cam disc can be used as shown 25 in Figures 34 and 35.

In the diagrammatically illustrated gearing in Figure 16, the central gears have a tooth count difference Ai= 2 The cam disc 70 carries a zig-zag shaped metal sheet as a planet gear.

Figure 17 A shows diagrammatically the mutual association of the four 30 structural components of the invention, namely, a pair of internally toothed central gears 42 and 44, an externally toothed planet gear 50 and the cam disc 70 In the table of Figure 17, there is specified which part ( 70 or 42 or 44) is driven, which part ( 42 or 44 or 50 or 70) is held stationary, and which part ( 42 or 44 or 50) can be connected to the driven shaft 35 The rotary direction is indicated by an arrow in the column "Drive" In the column "Reduction", there is found, in addition to the reduction ratio, the rotary direction of the driven shaft, also indicated by an arrow; if the latter arrow is in the same direction as the arrow in the column "Drive", then it is being driven in the same rotary sense; if the arrow in the column "Reduction" is opposite in direction 40 to the arrow in the column "Drive", then the driven shaft is being driven in a rotary sense opposite that in which the drive shaft is driving.

For the calculation of the reduction ratios in the column "Reduction", the following tooth counts were assumed:

central gear 42, A 42 = 80 45 central gear 44 i 44 = 78 planet gear 50 A 50 = 79.

Figure 18 is a view similar to that of the left-hand side of Figure 2; Figure 19 is a view of part of a zig-zag shaped bent metal sheet The individual upwardly projecting spikes of the sheetmetal, as viewed in Figure 19, form the teeth 48 of the 50 planet gear 50 (at the right-hand side of Figure 18), when the sheetmetal that is bent into the zig-zag shape is placed about the cam disc 70 The teeth 48 then engage in or mesh with the virtual tooth row, as do the teeth 88 of Figure 2 The direct placement of the zig-zag shaped, bent metal sheet on the cam disc 70 is possible with adjusting drives because, with the latter, due to the relative rotary motion of 55 the parts, friction that may occur is negligible For more rapid rotary motions, the structure shown in the left-hand side of Figure 18 is advantageously selectable.

Therein, the zig-zag shaped, bent metal sheet is seated on a ring 132 and forms therewith the planet gear 50 The ring 132 is mounted by means of roller bearings 69 72 on the cam disc 70 in a manner that, during movement of the ring 132 relative to 60 the cam disc 70, only minimal friction occurs.

A zig-zag shaped, bent metal sheet according to Figures 18 and 19 has the same advantage as the elastically and bendably constructed planet gear of the righti A 1 1 I s 1,567,331 hand half of Figure 2 When there is non-uniform division of the virtual tooth row, the zig-zag formed, bent metal sheet offers thereby not only the advantage of being manufacturable relatively easily and inexpensively, but also, of being able to equalize or compensate for the non-uniform division of the virtual tooth row.

In a modified construction, instead of one zig-zag-like bent metal sheet, at 5 least two are used lying one on top of the other and forming the teeth of the planet gear A tooth row of high elasticity and strength is obtained thereby The advantage of such a "multi-layered" zig-zag metal sheet is comparable to the advantage offered by a multiwire cable against a steel rod of the same crosssection.

A possibility of providing a shiftable disposition of the teeth on the planet gear 10 is shown in Figures 20, 21 and 22.

According to Figure 21 (a partial side elevational, partial sectional view) and Figure 22 (a top plan view), the planet gear 50 is provided with lateral guides 134 and 136 (for example, in the form of rings), which are overlapped by projections 138 and 140 of the teeth 48 The tooth 48 shown in perspective view in Figure 20 is 15 thereby shiftable in peripheral direction of the planet gear 50.

Figure 23 shows a section of a planet gear 50 which is formed with bores 142 that extend in axial direction through the entire planet gear Alternatingly, the bores 142 are open at the periphery thereof at one and the other side, respectively of the planet gear Such a planet gear can change its dimensions in peripheral 20 direction within given limits (such as 5 % for example) and thereby match or accommodate to the cam disc In the bores 142 of the planet gear that are open to the outside, a pair of resilient sheetmetal strips 144 having a substantially double-S cross-section are received in such manner as to form a spring in the shape of a figure-eight in cross-section that is divided, respectively, at the top and bottom 25 The lower section of the spring can pivot about a small angle in the bore 142 The upper section thereof carries a tooth 48, the flank part 168 of which is formed with an inner circular recess 146 which engages around the rounded upper part of the spring in a manner that the tooth 48 can pivot on the spring The recesses 146 have gripping traction to prevent the teeth from falling out 30 Such a mounted tooth has numerous degrees of freedom in the plane of the drawing of Figure 23:

1 The turning of the spring in the planet gear and of the tooth on the spring permits a variation of the angle between the tooth bisector lines 148 and the planet gear (the curved double-headed arrow 150 in the spring of the tooth on the left 35 hand side of Figure 23).

2 The yieldability of the upper part of the spring that engages in the tooth simultaneously permits a variation of the flank angle 2 a and the height or elevation h of the tooth; a desired dependence of the height h upon the variation or change of the flank angle 2 ac is adjustable by suitable dimensioning 40 The construction shown in Figure 23 thus permits the matching or accommodation of the tooth row of the planet gear to the virtual tooth row which is formed of the tooth rows of the internally toothed (or externally toothed) central gears.

Figure 24 shows a relatively short section of a planet gear 50 with a pivotable 45 bearing of a tooth 48 formed with a slot 178 extending in the region of the tooth bisector 148 The planet gear 50 has a concave cylindrical surface serving as a pivot bearing bed 152 of the tooth 48 If the center of curvature of the concave cylindrical surface 152 were identical with the point 154 of the tooth 48 (the rounding off thereof being ignored in this regard), then every swing of the tooth in 50 the directions of the double-headed arrow 50 would effect no change in the height of the tooth If the center of curvature of the concave cylindrical surface 152 should, however, be located outside the point 154, every swing of the tooth 48 then effects a change in the height thereof Through suitable selection of the curvature of the concave surface 152, any desired relationship between the change in the 55 height of the tooth and the swinging thereof in direction of the doubleheaded arrow 150 can be achieved Moreover, just as for Figure 23, there is produced obviously a dependence between the height of the tooth and the tooth flank angle thereof.

According to Figure 25, the planet gear 50 is formed with semicylindrical 60 recess 146 A cylindrical spring 156, which is open at a location 158 thereof, is received in the recess 146 A tooth 48 having two flank parts 168 that are connected at the top thereof, as viewed in Figure 25, is seated on the spring 156, just as it is seated on the spring metal sheets 144 in Figure 23 Such a construction is less costly than the bipartite, complex spring of Figure 23 It has the disadvantage similar to 65 1.567 331 1 2 that of the construction of Figure 23, however, that the tooth can fall out of the housing thereof This is avoided by the "tooth clip" 160 shown in Figure 26 which, laterally of the teeth of the central gears and around the entire planet gear, connects the teeth 48 of the planet gear to one another.

The left-hand part of Figure 2, as well as Figures 18 and 19, teaches that the 5 teeth of the planet gear can be found of a zig-zag shaped, bent metal sheet A given variability of the flank angle, a given variability of the height of the teeth, as well as a given mobility or shiftability in peripheral direction are thereby attained beforehand A disadvantage of such a simple zig-zag shaped metal sheet is that the 19 teeth are not completely independently shiftable from one another in peripheral 10 direction It is sufficient, however, that if only every second tooth of the plant gear is available, a construction according to Figure 27 is utilizable This is formed also -of a bent metal plate or sheet; a section of the metal sheet is bent triangularly to a tooth 48, the next succeeding section is bent into an arc 162 acting as an articulating joint, that is so low that it does not engage in the corresponding gap or 15 space of the virtual tooth row, the then following section is again bent into a tooth 48, the succeeding section again into an arc 162, and so forth Such a bent metal sheet formed as a tooth row of the planet gear, only every second (or possibly every third or fourth) tooth of which is present, and hinge-like arcs 162 being inserted 29 therebetween, is especially inexpensive to produce and adequately satisfies the 20 requirement for an independent mobility or shiftability of the individual teeth 48 in peripheral direction.

Quite generally, it should be noted that, for large diameters, not alf teeth of the planet gear must be present; it is sufficient, for example, if only every third tooth is present A considerable economy in the production thereof is thereby realizable 25 Figure 28 shows a section of a planet gear which is formed of individual guide shoes 164 that are held together by resilient cylinder pins 166 in such manner that they are capable of moving slightly toward one another in peripheral direction.

Every guide shoe 164 is formed with a radially outer concave surface 1following, 39 in cross-section, a circular line, the tooth 48 being seated on the respective concave 30 surface 128 The concave surface is curved in a manner that the center of curvature is disposed in the point of the tooth 48 If the tooth 48 accordingly shifts on the concave surface 128, the location of the point and the height thereof, accordingly, thus remain unchanged, only the orientation of the tooth 48 to the guide shoe 164 varies 35 The tooth 48 of Figure 28, like that of Figures 23 to 26, is formed of two flank parts 168 Figure 29 shows a tooth, both flank parts 168 of which are connected hinge-like to one another by means of a pin 188 disposed in the region of the tooth point.

Figure 30 shows a quarter of a gearing, in radial section, similar to that of 40 Figure 3 except that in Figure 30, however, the cam disc 70, the rollers 72 serving as force-transmitting bearings, the planet gear 50, the teeth 48 of the planet gear, as well as both central gears 42 and 44 are fully illustrated In the gearing shown in Figure 30, the tooth count difference of both central gears is four, and the M 40 reduction ratio 1:15 The high proportion of the illustrated teeth that are in mesh is 45 readily recognizable.

The planet gear 50 is variable in length in the peripheral direction thereof, because it is formed with bores 142 extending perpendicularly to the plane of the drawing of Figure 30 in a manner similar to that for the planet gear of Figure 23.

Moreover, the planet gear of Figure 30 is split at the interruption locations 170 in 50 | order to prevent the formation of internal stresses, for example, due to temperature variations The opened interruption location 170 of the planet gear is loosely held together by a strap lock 172, shown in broken lines; the strap lock 172 being constructed similarly to the lock of a bicycle chain except that it is resiliently flexible or yieldable in peripheral direction, however 55 During the introduction of force through the rollers 72, a bearing band 174, such as a steel belt, for example, is inserted between the rollers 72 and the planet gear 50 so that the rollers 72 do not penetrate into the bores 142 facing toward them This bearing band 174 is formed with a slot 176 so that the periphery thereof can vary (for example, with temperature variations) without arching or buckling 60 The slot 176 extends at an inclination across the bearing band 174, so that the rollers 72, which extend in direction of the central axis, do not penetrate into the inclined slot 176 (note Figure 31).

The individual teeth 48 of the planet gear in Figure 30 are mounted in the same manner as those described with respect to Figure 24 In this regard, the center of 65 I 1,567,331 14 1,567,331 14 the concave cylindrical surface 152 forming the pivot bearing bed is located exactly in the respective points 154 of the teeth, so that the height of the teeth do not change when the teeth pivot in the pivot bearing bed 152 Matching or accommodation of the direction of the teeth 48 to the direction of the tooth gaps or spaces of the virtual tooth row is thereby attained (if one considers that the term 5 "direction" of a tooth or of a tooth gap or space is that of the tooth bisector 148).

In addition, as in Figure 24, the teeth 48 of Figure 30 are formed, respectively, with a slot 178 What is achieved thereby is that the tooth flank angle matches the angle of the virtual tooth gap or space; if the tooth flank angle increases, the height of the teeth 48 then decreases, if the tooth is compressed, however, into a smaller 10 tooth flank angle, the tooth then becomes higher Through suitable dimensioning, adjustment to a desired relationship between both of these variations is able to be effected Figure 30 accordingly shows all of the degrees of freedom that may be given to the tooth 48 of the planet gear 50 so that it adjusts optimally to the virtual tooth row It is, moreover, also possible to make the teeth of the central gears 42 15 and 44 somewhat pivotable or swingable To this end, elongated recesses 180 are formed in the forward central gear 44 and permit connection of tooth foot 182 to the body 186 of the central gear 44 only through a narrow bridge 184 The bridge 184 acts as a joint about which the respective tooth of the central gear 44 can swing or pivot slightly The same feature is applicable to the teeth of the rearward central 20 gear 42 but is not illustrated in Figure 30 in the interest of maintaining the clarity thereof.

Both tooth rows of the central gears 42 and 44 are disposed in such a manner that those circles which bisect the teeth in height coincide in a circle 80 A consequence thereof is that the radii of the root circles 561 and 562 of the central 25 gears 42 and 44 are different The central gear 44 with the greater tooth count (lesser division) has the smaller root circle 562 During rotation of the cam disc 70 about the central axis 60, the teeth 48 of the planet gear are introduced with constant radial velocity into the virtual tooth gaps or spaces Due to the pivotability of the teeth 48 in the pivot bearing beds 152 and due to the slots 178 formed in the 30 teeth, an optimal matching to or accommodation of the virtual tooth row is achieved The elasticity of the planet gear 50 is also conducive to the compensation for or equilization of the varying division of the virtual tooth row.

The inner contour or outline of the planet gear 50 forms a kink or inflection point 98 At the corresponding location thereof (note Figure 3), the cam disc 70 has 35 a rounded-off section 100 which does not exactly correspond in form to the inner contour of the planet gear 50 This rounded-off section 100 prevents the rollers 72 from running over a tooth point, which would have resulted in the respective tooth having an insecure support.

The rearward central gear 42 has the greater root circle 561 The tooth gap 40 points thereof, which are shown in broken lines, thus extend (to the foot circle 561) farther outwardly than do the solid-line tooth gap points of the forward central gear 44 which extend only to the smaller root circle 562 The planet gear 50 rotates in direction of the arrow 112 i e counterclockwise as viewed in Figure 30, so that positions 190, 192, 194 and so forth are run through in succession The deeply 45 inserted tooth is located in the position 192 out of contact with the flanks of the rearward central gear 42 From the position 192 to the position 164, the tooth gap point of the rearward central gear 42 shifts through an amount A 9 p with respect to the solid-line tooth gap point of the forward central gear 44 In the position 194, the tooth 48 of the planet gear is disposed with the flank at the right-hand side thereof 50 at both flanks at the left-hand side of the central gears 42 and 44; in the position 190, the tooth 48 is disposed with the flank at the left-hand side thereof at both flanks at the right-hand side of the central gears 42 and 44 During the movement through these three positions in direction of the arrow 112, the tooth 48 thus changes from the tooth flank at the right-hand side to the tooth flank at the left 55 hand side of the rearward central gear 42 and, in the intermediate position 192, it is in meshing engagement only with one central gear, namely the forward central gear 44 which has the smaller root circle, and in fact owing to the roundingoff section co-operating with the different root circles The course of movement during the reversal of the direction of movement of the tooth 48 is thereby not disrupted 60 Due to the central gear 44, the tooth 48 remains prestressed and passes virtually free from losses again to the central gear 42 when the spring or resilient energy of the tooth 48 has run out.

Figure 31 is a top plan view of a section of the bearing band 174 that is formed with the slot 176 aforementioned with respect to Figure 30 65 Figure 32 shows in a longitudinal sectional view the gearing according to the invention Figure 1 shows a cross-sectional view along the line I-I in Figure 32.

Figure 33 shows a further embodiment in a view that could have been taken along the line I-I in Figure 32 i e a view corresponding to that of Figure 1.

^ 5 Figure 33 shows a true-to-scale radial cross-sectional view of a gearing 5 according to the invention having a reduction ratio of ten; in a singlestage construction, such a small reduction ratio is practically the realizable critical or borderline case Just as in Figure 30, the root circles 561 and 562 of the central gears 42 and 44 are different The virtual tooth gap root curve 64 is displaced or dislocated by an eccentricity E with respect to the center 58 The planet gear 50 is 10 slotted at the interruption location 170 and is per se covered with a slide layer 196 at the inside thereof The teeth 48 are pivotally mounted in the planet gear in a manner similar to that in Figure 30 They can therefore be accommodated exactly to the virtual tooth row.

Also, in this embodiment, it is sufficient if both central gears 42 and 44 are 15 mounted in the planet gear; it is unnecessary to mount one central gear in the other central gear The cam disc 70 is advantageously driven through an elastic coupling.

A gearing system such as is shown in Figures 34 and 35 has proven to be practical Figure 34 is an axial cross-sectional view of Figure 35 taken along the line XXXIV-XXXIV in the latter, and Figure 35 is a radial sectional view of Figure 34 20 taken along the line XXXV-XXXV therein.

The closely hatched cam disc 70 of Figure 35 is formed of two halves 703 and 704 and is provided with four coupling bores 198 wherein respective coupling pins or bolts 200 are received An end of the respective bolts 200 projects into a suitable 2 W 5 -recessed flange 202 of the drive shaft 76 (note Figure 34) 25 The cam disc 70 carries the bearing band 174 on rollers or needle bearings 72.

The bearing band 174 is formed with an inclined slot 176 (note Figures 30 and 31).

The bearing band 174 is surrounded by the planet gear 50, the teeth of which mesh with the teeth of both central gears 42 and 44 The central gear 42 at the left-hand side of Figure 34 is mounted on a bearing 204 in the central gear 44 at the right 30 hand side of that figure; the central gear 42 being rigidly connected to the drive shaft 78 Seals 206 and 208 provide for the sealing of the gearing from the outside (note Figure 34).

The tooth count difference of both internally toothed central gears 42 and 44 is four In order to avoid the obstruction of details in Figure 35, the planet gear and 35 the teeth thereof, which are adjustable in accordance with the previously described figures, are not illustrated in Figure 35 Only the virtual tooth gap root curve 64 is shown in (heavy) dot-dash lines The curve 64 is similar to the curve path 641 of Figure 13 With regard to the rounded-off portion 100 of Figure 3, the cam disc 70 according to Figure 35 can be produced from two semi-circular halves 703 and 704 40 The spacing between both semi-circular halves 703 and 704 is adjustable by two screws 210, which are shown in plan view in Figure 34 and in diagrammatic longitudinal sectional view in Figure 35 A fitting hole 212 visible only in Figure 34 is located between both screws 210 for receiving therein a fitting pin through both halves of the cam disc 70, respectively above and below, in order to adjust both 45 halves accurately The screws 210 afford an exact adjustment of the spacing between both halves of the cam disc during assembly of the gearing; no high precision need be maintained during manufacture, accordingly, because the required accuracy of adjustment during assembly can be achieved by the adjustment of the 1 50 screws 210 During subsequent wear the affected parts can be afteradjusted by 50 means of the screws 210 Both halves of the cam disc 70 are guided toward one another at locations 214 by groove and spring.

Between the drive shaft 76 and the cam disc 70, only a force-locking, and no form-locking coupling, however, ought to be provided For this reason, an elastic coupling is provided and, in fact, in such a manner that the coupling pins 200 carry 55 thick rubber rings or sleeves 216, which couple the pins 200 elastically to the coupling bores 198 Other elastic couplings or tooth couplings are contemplatable.

Figure 36 shows a control gearing in a longitudinal sectional view A first cam disc 701 is driven by the drive shaft 76 and is rigidly connected to a second cam disc 702 disposed offset by an angle of 180 therefrom The cam disc 702 is mounted in 60 the bearing 218 in the driven shaft 78.

Every cam disc 701 and 702 carries, on rollers 72, a slotted bearing band 174 as well as a planet gear 501 or 502 with teeth which can match or accommodate the respective virtual tooth row, in accordance with the foregoing representations.

Three central gears 421, 422 and 423 are all mounted within one another in two 65 1 i 1 i 1 I 1,567,331 is 16 1,567,331 16 layers 204 The first central gear 421 facing the drive shaft 76 meshes with a tooth row thereof in the tooth row of the planet gear 501 of the first cam disc 701 The middle central gear 422 has two tooth rows, the first of which meshes with the teeth of the planet gear 501 of the first cam disc 701, and the second of which meshes with the teeth of the planet gear 502 of the second cam disc 702 The teeth of the 5 third central gear 423 facing toward the driven shaft 78 mesh with the teeth of the planet gear 502 of the second cam disc 702.

Altogether, five engageable and disengageable couplings Kl, K 2, K 3, K 4 and K 5 are provided:

1 Through the coupling Kl, a coupling ring 222 coaxially surrounding the 10 drive shaft 76 can be coupled to the housing 220 The central gear 421 surrounds an inner section of the coupling ring 222; between both thereof, a coupling K 2 is provided.

2 The driven shaft 78 is rigidly connected to a sleeve or bushing 224 which surrounds the three central gears 421, 422 and 423 Through a respective coupling 15 K 3, K 4, or K 5, the sleeve or bushing 224 can be coupled to one of the central gears, respectively.

In order to attain a reduction ratio between the drive shaft 76 and the driven shaft 78, at least two couplings must be engaged As noted hereinbefore, two different possibilities of effecting reversal are realizable Various proposals with 20 respect to the numbers of teeth of the central gears 421, 422, and 423 have also been presented.

Various combinations of fixed and loose clutches or couplings can be derived from the following table:

Fixed Couplings Loose Couplings Rotary Speeds Taken Off 25 Kl K 2 respective rotary speeds at K 3, K 4 or K 5 K 2 +K 3 Kl i=l, no reduction Kl+K 2 K 3 respective rotary speeds at K 4 or K 5, the rotary speed 30 of which is opposite to the driveshaft 76 i e two reversals The following advantages over heretofore known planetary gearing are achieved by the planetary gearing of the invention: 35 1) The planetary gearing according to the invention can transmit considerably higher torque than prior-art gearing of the same dimensions and weight.

2) The resultant force action upon each individual tooth of the planet gear is disposed perpendicularly to the periphery of the planet gear The result thereof is:

a) that the individual teeth of the planet gear are independent of one another, 40 b) that the planet gear is not stressed in torque or bending load, c) that the planet gear can be yieldably or resiliently constructed without having to be supported by the cam disc and without having to be held in the form wherein it engages in the virtual tooth row, in accordance with the invention, contrarily, the virtual tooth row determines the form of the 45 planet gear or, if the planet gear is yieldable or resilient, the virtual tooth row impresses the form of the planet gear.

3) The planetary gearing according to the invention is not reversible but is rather self-locking Heretofore known gearing were not reversible due to selflocking i e it had a high tooth friction and bearing friction and accordingly a poor 50 efficiency; the gearing according to the invention is not reversible due to the kinematic principle thereof (note hereinbelow the description respecting Figure 4) and simultaneously possesses a high efficiency.

4) The planetary gearing according to the invention operates without play between the tooth flanks of the teeth of the internally toothed (or externally 55 toothed) central gears, on the one hand, and of the planet gear, on the other hand, and with uniform rotary speeds (without angular acceleration).

5) The planetary gearing is suited in single-stage construction for gear reductions of between substantially 10 and 300.

6) The individual parts of the planetary gearing according to the invention 60 center one another i e they are "self-centering".

7) All of the foregoing advantages are attainable with relatively simply constructed and relatively simply manufacturable planetary gearing.

17 1,567,331 17

Claims

WHAT WE CLAIM IS:-

1 A planetary gearing comprising a toothed planet gear, first and second relatively rotatable central gears having juxtaposed rows of teeth meshing with said toothed planet gear, said juxtaposed rows of teeth having different numbers of $ teeth and forming at least two virtual tooth rows by overlapping, each tooth gap of 5 each virtual tooth row being formed by a side of one of the teeth of the first central gear and a side of one of the teeth of the second central gear and each virtual tooth row defining a curve passing through the roots of the tooth gaps of that virtual tooth row, and a cam rotatable about an axis and having a cam outline engaging the planet gear for guiding and driving the planet gear, all of the teeth being of 10 substantially triangular cross-section and having substantially flat sides, the sides of the teeth of the planet gear substantially flatly engaging the sides of the teeth of one of the virtual tooth rows, the pitch of each meshing planet gear tooth being substantially equal to the pitch of the respective meshing tooth of said virtual tooth row, and the centroid of the area enclosed by the cam outline coinciding with the 15 centroid of the area enclosed by the curve passing through the roots of the tooth gaps of said one virtual tooth row.

2 A planetary gearing as claimed in Claim 1, wherein the number of teeth of the planet gear which are in force transmitting engagement with said one virtual tooth row is substantially equal to the total number of teeth of the planet gear less 20 the difference between the numbers of teeth of the pair of central gears.

3 A planetary gearing as claimed in Claim 1 or Claim 2, wherein said central gears are internally toothed and said planet gear is externally toothed.

4 A planetary gearing as claimed in Claim 1 or Claim 2, wherein said central gears are externally toothed, said planet gear is internally toothed and surrounds 25 said central gears, and said cam is operatively connected to the planet gear from the outside thereof.

A planetary gearing as claimed in any one of Claims 1 to 3 including at least another internally toothed central gear meshing with said toothed planet gear.

6 Planetary gearing according to Claim 5, wherein all of said internally 30 toothed central gears are coaxially disposed one behind the other, and said planet gear is externally toothed.

7 Planetary gearing according to Claim 5 or Claim 6, wherein said pair of central gears have the same number of teeth, respectively, and are disposed on both sides of a central gear having a different number of teeth 35 8 Planetary gearing according to Claim 1, Claim 2 or Claim 4 including at least another externally toothed central gear meshing with said toothed planet gear.

9 Planetary gearing according to Claim 7, wherein all of said externally toothed central gears are disposed one behind the other and said planet gear is 49 internally toothed and surrounds the central gears, and said cam is operatively 40 connected with said planet gear from the outside thereof.

Planetary gearing according to Claim 8 or Claim 9, wherein said pair of central gears have the same number of teeth, respectively, and are disposed on both sides of a central gear having a different number of teeth.

11 Planetary gearing according to any preceding claim, wherein the number 45 of teeth of said planet gear lies between the respective numbers of teeth of said central gears.

12 Planetary gearing according to any preceding claim, wherein the difference in the numbers of teeth of said central gears is between one and six.

SO 13 Planetary gearing according to any preceding claim, wherein the difference 50 in the numbers of teeth of said central gears is two, and said curve passing through the roots of the tooth gaps of said one virtual tooth row is substantially a circle having a center disposed eccentrically to the central axis of said central gears by substantially half the height of a tooth of said planet gear.

14 Planetary gearing according to any preceding claim, wherein the difference 55 in the numbers of teeth of said central gears is two, and wherein circles halving the teeth of said central gears in the height thereof coincide in a common circle so as to form a virtual tooth row having constant pitch.

Planetary gearing according to Claim 1, wherein said cam has a contour substantially similar, in mathematical sense, to said curve, locations thereon, at 60 which directions of tangents thereto vary irregularly, having rounded portions, and including bearing means disposed between said curve and said contour of said planet gear for transmitting the teeth as well as the force thereof, and means for adjusting the flank angle of said teeth to said virtual tooth row so as to compensate for non-uniform pitch thereof As 1 v 16 Planetary gearing according to Claim 15, wherein said bearing means comprise rollers.

17 Planetary gearing according to Claim 15 or Claim 16, wherein said flank angle adjusting means comprises means for independently shifting the teeth of said planet gear in peripheral direction 5 18 Planetary gearing according to Claim 15 or Claim 16, wherein said flank angle adjusting means comprises means for independently shifting the teeth of said planet gear in radial direction.

19 Planetary gearing according to Claim 15 or Claim 16, wherein said flank angle adjusting means comprises means rendering said teeth of said planet gear 10 elastically deformable.

Planetary gearing according to any of Claims 15 to 19, wherein said pair of central gears are internally toothed and said planet gear is externally toothed, and one of said central gears having the greater number of teeth has a root circle connecting the roots of the teeth which is the smaller of the respective root circles 15 for the pair of central gears.

21 Planetary gearing according to any of Claims 15 to 19, wherein said pair of central gears are externally toothed and said planet gear is internally toothed, and one of said central gears having the greater number of teeth has a root circle connecting the roots of the teeth which is the larger of the respective root circles 20 for the pair of central gears.

22 Planetary gearing according to any of Claims 15 to 21, wherein said teeth of said planet gear are variable in height and are spring-loaded in direction toward said curve.

23 Planetary gearing according to Claim 22, wherein said tooth flanks are 25 uneven.

24 Planetary gearing according to any of Claims 15 to 23, wherein the tooth flanks of said central gears are flat, said curve being composed of circular arcs about respective centers and being calculable from an equation of the following type: 30 rv =r sina^; ( -As wherein:

rv=the distance of one tooth gap root (at a location m) of the virtual tooth row from the central axis of the central gears, r=the radius of the root circle of the central gears with respect to the central 35 axis of the central gears, A 9 = the angular difference (peripheral spacing) of the adjacent tooth gap roots of the central gears at the location m calculated from the equation:

( < -o vc) = 36 60 ( 1 z_) m ()) where 40 aa=half the flank angle of the tooth gaps of the used virtual tooth row, a=half the flank angle of the teeth of the planet gear, (TV-aam=half the difference between the flank angle of the used virtual tooth row, on the one hand, and the flank angle of the plane gear, on the other hand, at the location m, 45 A,=the number of teeth of a first central gear, A 2 =the number of teeth of a second central gear, and m=the number (ordinal) of the tooth under consideration, as counted from a location at which av-a= O; and from the equation: 50 (p-va i (pa= A a=A( 3) where q=the angular distance of the vertical tooth bisector of one tooth of the planet gear from the location m= O; and I 1,567,331 ip=the angular distance of the vertical tooth bisector of the corresponding tooth of the virtual tooth row from the location m= 0:

2 a,=the flank angle of the tooth gap of the virtual tooth row at the location m according to the foregoing equation ( 1), and As=the spacing of straight connecting lines between the tooth gap points called 5 AT of the central gears from the root circle.

Planetary gearing according to Claim 24, wherein the center of mass of the centers of said circular arcs is located on the central axis of the central gears, said cam being formed of two relatively adjustable halves with a circular contour, the contour of each half extending respectively over somewhat less than a half circular 10 arc having a center disposed in a center of mass of that pair of centers of said circular arcs, which compose said curve, that is more distant from the respective contour.

26 Planetary gearing according to Claim 15, wherein each tooth is formed of two flank parts articulatingly connected at the point of the tooth, said flank parts 15 having therebetween a spring element tending to bias said flank parts away from one another, said flank parts having respective free ends spaced from said planet gear.

27 Planetary gearing according to Claim 15, wherein said planet gear is i 2 formed with bores extending in axial direction and being alternatingly opened to 20 the outer and the inner side of said planet gear, respectively, a pair of double-S shaped spring plates being received in the bores opened to the outside and forming a figure eight-shaped spring having a radially inner section thereof disposed in said bore and a radially outer section thereof pivotally received in substantially circular recesses formed in said flank parts 25 28 Planetary gearing according to Claim 15, wherein every individual tooth is seated on a guide shoe, and the guide shoes being resiliently connected together to said planet gear.

29 Planetary gearing according to Claim 15, wherein said planet gear is formed with a break in the continuity thereof, and including a peripherally resilient 30 strap lock bridging said break.

Planetary gearing according to Claim 25, wherein rollers serving as forcetransmitting bearings are provided between said cam and said planet gear, and a bearing band is inserted between said rollers and said planet gear.

31 Planetary gearing according to Claim 30, wherein said bearing band is 35 formed with a slot inclined to the longitudinal direction of said bearing band.

32 Planetary gearing according to Claim 15, wherein the tooth feet of said central gears are connected only by a narrow, bendable bridge to the body of the respective central gear.

j 33 Planetary gearing according to Claim 15, wherein said teeth of said central 40 gears are formed with cooling bores.

34 Planetary gearing according to Claim 15, wherein said planet gear comprises a metal sheet having a zig-zag formed cross-section disposed about the periphery of said cam and defining the teeth of said planet gear.

35 Planetary gearing according to Claim 34, wherein said zig-zag formed 45 metal sheet has a plurality of layers.

36 Planetary gearing according to Claim 34 including a ring member whereon said zig-zag formed metal sheet is mounted, and anti-friction bearing means supporting said ring on said planet gear.

i O 37 Planetary gearing according to Claim 15, wherein said cam is mounted on 50 bearing means therefor in said planet gear which, in turn is mounted in said virtual row only force-lockingly, said bearing means in said planet gear being the sole means for mounting said cam.

38 Planetary gearing according to Claim 1 or Claim 2, including at least another cam mounted coaxially with said first-mentioned cam on a common drive 55 shaft, each of said cams carrying a respective planet gear, at least three internally toothed central gears being couplable through engageable clutches with said common drive shaft, a first of said central gears facing toward said drive shaft having teeth meshing with teeth of a second of said central gears, which also faces toward said drive shaft, and including a middle central gear having two rows of 60 teeth, one of said rows of teeth meshing with the teeth of a planet gear adjacent to said drive shaft and the other of said rows of teeth meshing with the teeth of the other of said planet gears, said other of said planet gears facing toward a driven shaft.

39 Planetary gearing according to Claim 38, wherein said first central gear 65 I 1,567,331 facing toward said drive shaft is couplable by a releasable clutch to a housing for the planetary gearing.

Planetary gearing according to Claim 39 including a clutch ring mounted on said drive shaft between said housing and said first central gear, said clutch ring being couplable through a releasable clutch with said first central gear and through 5 a further releasable clutch with said housing.

41 Planetary gearing according to Claim 40, wherein a clutch between said first central gear and said driven shaft is releasable while said clutch between said first central gear, said clutch ring and said housing is fixed so that selectively through the clutch of said third central gear and said clutch of said middle central 10 gear, a rotation can be taken off having a rotary sense opposite that of said drive shaft.

42 Planetary gearing according to any of Claims 38 to 41, wherein said third central gear has a number of teeth equal to the number of teeth of one of the rows of said middle central gear, the other of the rows of said middle central gear having 15 a number of teeth equal to the number of teeth of said first central gear, and the number of teeth of said first central gears are unequal.

43 Planetary gearing according to any of Claims 37 to 40, wherein the said number of teeth of one of the rows of said middle central gear is equal to the number of teeth of the other of the rows of said middle central gear, the numbers of 20 teeth of said first and third central gears being different from one another and from the numbers of teeth of both said rows of said middle central gear.

44 Planetary gearing according to any of Claims 38 to 41, wherein the numbers of teeth of said first and third central gears as well as both said rows of said middle central gear are dissimilar 25 Planetary gearing according to any of Claims 38 to 45, wherein the largest diameter of one of said cams is offset with respect to the largest diameter of the other of said cams.

46 Planetary gearing according to Claim 45, wherein the extent of offset between the largest diameters of both said cams is 1800 30 47 Planetary gearing according to Claim 38, wherein more than two cams are mounted on a common drive shaft, wherein the number of central gears is greater by one than the number of cams, and those central gears disposed between two other central gears carry two rows of teeth of which one row of teeth meshes with one planet gear and, the other row of teeth with an adjacent planet gear 35 48 Planetary gearing substantially as hereinbefore described with reference to and as shown in Figures 1 and 32 or modified as shown in Figures 3 to 7 or modified as shown in Figure 16 or modified as shown in Figure 17 A or modified as shown in Figure 33 of the accompanying drawings.

49 Planetary gearing substantially as hereinbefore described with reference to 40 and as shown in Figure 2 or modified as shown in Figures 18 and 19 or modified as shown in Figures 20 to 22 of the accompanying drawings.

Planetary gearing substantially as hereinbefore described with reference to and shown in Figures 30 and 31 of the accompanying drawings.

51 Planetary gearing substantially as hereinbefore described with reference to 45 and as shown in Figures 34 and 35 of the accompanying drawings.

52 Planetary gearing substantially as hereinbefore described with reference to and as shown in Figure 36 of the accompanying drawings.

53 Planetary gearing as claimed in Claims 47 to 51 having planet gear teeth substantially as hereinbefore described with reference to and as shown in one of 50 Figures 23 to 29 of the accompanying drawings.

BALCKE-DURR AKTIENGESELLSCHAFT, Per: Boult, Wade & Tennant, 34 Cursitor Street, London, EC 4 A 1 PQ, Chartered Patent Agents.

Printed for Her Majesty's Stationery office, by the Courier Press, Leamington Spa 1980 Published by The Patent Office, 25 Southampton Buildings, London WC 2 A l AY, from which copies may be obtained.

I 1,567,331